1. Field of the Invention
This invention relates generally to servo valves for operating hydraulic actuators, and, more specifically, to those servovalves used in safety-critical applications or in force control applications or in acceleration control applications.
2. Background
Servo mechanisms are used to move aircraft control surfaces, to position machine tools, to move robotic manipulators, to simulate earthquakes, to test noise, vibration, to test harshness characteristics of vehicles, as energy sources for geophysical exploration, to grind eyeglass lenses, to move flight simulator cabins, to provide tactile feedback to joy sticks, to control pavement breaking machines, and for many other applications. Precision servo hydraulic actuators are typically controlled by servovalves. Non-precision actuators are typically controlled by similar but less expensive Proportional Valves. The following discussion will specifically address servovalves, but may apply equally well to proportional valves. Hereinafter, the term “servovalve” should be taken to include the so-called proportional valves.
In a typical form, a servovalve includes a spool valve which receives a flow of fluid from a source through one or more pressurized inlet ports, and whose position along its axis is controlled by variable volumes of fluid in two differential control chambers. These control chambers receive pressurized fluid from a prior stage of the servovalve. A difference in hydraulic force (defined as P×A, where P=Pressure and A=Area) between the two control chambers tends to accelerate the spool toward the chamber with lower hydraulic force. Two differential outlet ports are provided in a four-way valve for delivering pressurized fluid to either one of the opposite chambers of a linear actuator that includes a bidirectional piston mounted in a dual-chamber cylinder. The valve further includes at least one fluid return port that communicates with a fluid reservoir.
A conventional spool valve includes a spool having spaced-apart lands which is mounted for linear motion in opposite directions within a bore in a valve body. With the spool shifted to a first position, one or more pressurized fluid inlet ports are placed into hydraulic communication through a first outlet port with a first chamber of the actuator. Simultaneously, one or more low pressure fluid return ports are placed into hydraulic communication through a second outlet port with the opposite chamber of the actuator, thereby tending to move the actuator piston in one direction to change the position of a load which may be mechanically coupled to the actuator. Commonly, the instantaneous displacements of both the spool and the load are monitored by a Linear Variable Differential Transformer (LVDT) of any well-known type, or a Hall-effect position transducer, or a magnetostrictive transducer.
The actuator piston is urged to move in an opposite direction by shifting the servovalve spool in the opposite direction whereby one or more pressurized fluid inlet ports are placed into hydraulic communication through the second outlet port with the second chamber of the actuator. Simultaneously, one or more low pressure fluid return ports are placed into hydraulic communication through the first outlet port with the first chamber of the actuator, tending to move the actuator piston and connected load in the opposite direction.
The control law used in conventional flow control servovalves is to ideally make hydraulic fluid flow rate proportional to an input signal. The polarity of the input signal determines the direction in which the servovalve spool will move, while the magnitude of the input signal determines the velocity and displacement of spool movement. The magnitude of spool displacement from the center “null” position determines the magnitude of hydraulic fluid flow through the outlet ports. In the ideal case, which ignores flow restrictions, loading effects and variations in differential pressure across spool lands, flow from the outlet ports is proportional to the input signal.
The servo system may employ two stages in which the first stage is a torque motor which in turn controls the positioning of the valve spool of a second stage, which directly controls the actuator. To produce or control large dynamic loads, a third stage of amplification may be added as is well known. Thus in a multistage servovalve, the output stage is driven directly by one or more previous stages.
By Newton's third law, the actuator applies force to the load and the load applies equal and opposite force to the actuator. In order to operate on, e.g., move or rotate, a given load, the actuator force must be greater than the load force. The actuator force may be directly calculated by multiplying the hydraulic pressure on each side of the actuator piston by the piston area exposed to that pressure, and taking the difference of the two products, i.e., Actuator Force=(Pressure 1×Area 1)−(Pressure 2×Area 2) where Pressure 1 and Area 1 correspond to the pressure applied to and the area of a first side of the actuator piston and Pressure 2 and Area 2 correspond to the pressure applied to and the area of a second side of the actuator piston. For a typical control problem this equation may be simplified to: Actuator Force is proportional to (Pressure1−Pressure2)×Area Ratio, where the Area Ratio is the ratio of Area 1 to Area 2. Thus, this is the differential pressure times the piston area ratio. When the piston area exposed to hydraulic pressure is the same on both sides, the Area Ratio is 1. Otherwise, the Actuator Force is asymmetric, resulting in unequal force gain on oppositely directed strokes.
Since the hydraulic pressure on each side of an actuator piston is in direct fluid communication with its respective output port of a servovalve, actuator pressure may be measured either at the actuator or at the servovalve. The measurement site may be selected for convenience.
In some applications, whether the actuator piston Area Ratio is 1 or another value, force asymmetry might be caused by an asymmetric loading of the actuator. For example, an actuator which is oriented to move in a vertical axis and which supports a heavy load has load asymmetry caused by the force of gravity on the load. The gravitational force on the load increases pressure on the bottom side of the actuator piston in the quiescent state when the load is supported by hydraulic fluid. The quiescent state may be envisioned as a hydraulic lift which supports a load. An example of this situation is a machine which supports and shakes an automobile for noise, vibration and harshness (NVH) testing.
In other applications, force asymmetry may occur due to the actuator driving a nonlinear load, such as where the load is a nonlinearly compressible material, such as limestone.
The above described asymmetries tend to cause undesirable nonlinear distortion in an actuator's dynamic output. They tend to cause significant even-order harmonic distortion, and also odd order harmonic distortion. There is thus a need for a servo control system capable of compensating for force asymmetries related to a force servo actuator.
Servovalves typically have an initial transduction stage which changes a digital or analog electrical signal into a hydraulic signal. This transduction stage it typically followed by one or two stages of hydraulic amplification. Including the transduction stage, these are referred in the art as two-stage and three-stage servovalves. Proportional valves may have similarly configured stages and perform a function similar to servovalves, typically at lower cost and lower performance. Hereinafter, reference to servovalves should be understood to also include proportional valves.
For example, a servovalve hydraulic amplification stage might have a gain of 40 if its output flow rate is 40 times its input (command) flow rate. At the output ports A and B of a typical two-stage or three-stage four-way servovalve, pressurized fluid flows out port A at a commanded rate while fluid simultaneously flows into port B at a similar rate. Both fluid flows can also be simultaneously blocked, or reversed in direction. Fluid which the servovalve outputs is supplied by a pump, and fluid which it inputs goes typically to a reservoir from which the fluid subsequently returns to the pump. Filters, coolers, pulsation dampeners and numerous other devices well known in the art may be included in the hydraulic system.
A servovalve controls the rate at which it outputs and inputs fluid by controlling the position of its spool in a sleeve or bore. The spool is a device which moves along its central axis to control the opening and closing of four variable orifices. The spool has a center null position in which no fluid flows except for leakage. It can move one direction or the other from the null position in order to output fluid from output port A or B at a flow rate determined by its displacement from the null position.
In order to accurately control flow rate, a servovalve must accurately monitor and control the position of its output spool. In a two-stage servovalve, the monitoring function is typically done by a spring wire. In a three-stage servovalve, the monitoring function is typically done by a linear variable differential transformer (LVDT). Other well known types of position monitoring devices are sometimes used instead.
Sometimes a monitoring spring wire breaks, and sometimes an LVDT electrical cable develops a fault. If the monitoring device becomes inoperable for any reason, the servovalve cannot control the output spool position. The typical result of such a failure is extreme and erratic changes in output spool position resulting in maximal and unpredictable effects on the load. This sort of failure can sometimes cause catastrophic results.
Hydraulic pressure feedback serves as an auxiliary servovalve output spool position feedback in addition to its other functions. When the output spool is not in the null position, pressure feedback tends to push it back to null. Differential hydraulic pressure feedback (PFB) provides a restoring force, a negative feedback, and thus prevents erratic servovalve behavior in case the primary valve spool position feedback subsystem fails. A servovalve with PFB and with a failed primary valve spool position feedback subsystem tends to return its valve spool to null, gracefully stopping the actuator wherever it may be and preventing erratic and catastrophic motions.
The FAA has reportedly mandated that all early production Boeing 737 aircraft be modified to prevent an “uncommanded roll” failure reported by the Seattle Times to be the result of a servovalve feedback failure. Boeing's approach to the problem is to replace single-rudder tails with dual-rudder tails for redundancy. The same newspaper reported an incident of a Boeing 747 elevator failure, possibly caused by the same sort of servovalve failure. A seismic vibrator owned by a multi-national oil company was reported to have separated the top of the passenger cabin from the vehicle during violent shaking caused by a servovalve spool position feedback failure. While these incidents may be rare, they do happen, and the potential catastrophic results merit considering possible solutions.
3. Discussion of Related Art
P. F. Hayner, in U.S. Pat. No. 3,260,273, issued Jul. 12, 1966, entitled Motor Valve having Differential Pressure Feedback, teaches a pressure control hydraulic servovalve wherein a pilot valve positions a control valve member to control the application of fluid under pressure through an outlet in the control valve to provide an output differential pressure across an output actuator device. The differential pressure developed across the output actuator is coupled back to a movable pilot valve member to produce a substantially constant pressure across the actuator device.
U.S. Pat. No. 3,479,925, issued Nov. 25, 1969 to P. F. Hayner and D. G. Eldridge, entitled Hydraulic Signal and Summing System, discloses a pneumatic or hydraulic control system for controlling fluid pressure and flow to an actuator so that the actuator is controlled in accordance with a predetermined function of a plurality of fluid pressure signals. The fluid pressure signals are converted to mechanical forces which are applied to the mechanical structure which directly controls fluid flow from two or more nozzles producing a pressure differential representative of the predetermined function. This pressure differential is employed to meter the fluid pressure and/or flow to the actuator.
U.S. Pat. No. 4,372,193 issued Feb. 8, 1983 to L. R. Hall, entitled System with Constant Force Actuator, teaches a method of maintaining a constant force on an actuator by controlling a pilot differential pressure on the two ends of a pilot operated control valve. The technique passes excess flow to a downstream control valve.
U.S. Pat. No. 4,537,077 issued Aug. 26, 1985 to A. J. Clark and D. N. Maue, entitled Load Dynamics Compensation Circuit for Servohydraulic Control Systems, teaches an electronic compensation circuit which compensates for disturbance factors resulting from forces exerted by a specimen on an actuator. The compensation signal is an anticipatory signal compensating for the load dynamics of a test article on a vibratory test machine.
U.S. Pat. No. 5,128,908 issued Jul. 7, 1992 to D. K. Reust, entitled Pressure Feedback Servovalve for a Seismic Vibrator, teaches a method for reducing harmonic distortion associated with a hydraulic seismic vibrator apparatus using symmetric pressure feedback control. The patent teaches a method for converting the third stage of a three stage servovalve into a pressure control servovalve. The servovalve is converted into a pressure control valve by porting differential negative pressure feedback from the hydraulic output ports of the main stage. The pressure feedback is a differential flow of hydraulic fluid through two passageways based upon the differential pressure applied to the piston of the vibrator actuator. The differential pressure applied to the piston represents the load on the servovalve. The amount of feedback applied is determined by orifices in the feedback passageways. Hydraulic damping of the load is achieved by providing a restricted hydraulic path between the two output ports of the servovalve. The amount of damping is determined by an orifice.
The apparatus of the '908 patent is designed specifically for use on a seismic vibrator as used in geophysical exploration. Although this reference teaches structure for providing differential negative pressure feedback, it does not teach means for accommodating known or postulated force asymmetries related to the load actuator. More specifically, it does not teach or suggest providing hydraulic fluid feedback using fluid passageways in an intermediate land of the spool.
U.S. Pat. No. 5,230,272 issued Jul. 27, 1993 to J. Schmitz entitled Hydraulic Positioning Drive with Pressure and Position Feedback Control, discloses a hydraulic drive actuated by a CNC means. The feedback to an electronic servo amplifier may be switched to actuator position or actuator pressure.
U.S. Pat. No. 5,522,301 issued Jun. 4, 1996 to J. E. Roth, et. al. entitled Pressure Control Valve for a Hydraulic Actuator, describes a servovalve assembly for accommodating asymmetrical loading characteristic of unequal area pistons. This patent teaches use of two three-way servovalves, one valve being coupled to the load line of each chamber of the actuator. Asymmetrical actuator loading is compensated for by use of field-replaceable cartridges 82A and 82B characterized by a sleeve portion 220 and a spool 222 (FIGS. 4A, 4B and 6 of the '301 patent). The cartridges are installed in the respective servo valves with the spool portion of the cartridge in engagement with a land on the end of the spool of the servovalve. The differential force applied to the two cartridges multiplied by the ratio of their respective end areas dictates the force applied to the actuator piston. The force is a function of the surface area of one end of the cartridge spool and the pressure in the control passageway. Therefore, by increasing (or decreasing) the diameter of the cartridge spool portion the force applied by the cartridge to the spool and the force on the piston increases (or decreases). (See col. 9, lines 1-21). Restrictions 138A and 138B (FIG. 4A) provide damping.
U.S. Pat. No. 6,629,733, issued Aug. 7, 2001, to Dennis K. Reust, entitled Force Servo Actuator with Asymmetrical Nonlinear Differential Hydraulic Force Feedback, the entire contents of which are herein incorporated by reference, teaches a hydraulic force servo system which provides an apparatus for compensating for unequal loading forces applied to an actuator piston, and for unequal areas on opposite faces of the piston. Asymmetric nonlinear differential hydraulic force feedback from the output side of a force servovalve is summed with the hydraulic control signal inputs to the output stage of the servovalve at hydraulic summing junctions. Impedances offered by orifices in the feedback lines determine the amount of feedback. Nonlinear characteristics of the feedback method serve to compensate for nonlinearities in the servo actuator and system. The impedance ratio of the orifices is selected as a function of the known or postulated asymmetry of the loading forces to be applied to the actuator piston, and the ratio of the areas of the opposing faces of the actuator piston. Hydraulic damping further improves linearity and stability.
Notwithstanding the related art, there remains a need for a means to inexpensively create fail-safe servovalves, and further decrease the delay in pressure feedback control loops, and reduce cavitations inside servovalve bodies.